Variable valve gear device of internal combustion engine

ABSTRACT

A variable valve operating system of an internal combustion engine comprises a first variable mechanism  1  capable of variably controlling a lift characteristic of an intake valve  12 , and a second variable mechanism  2  capable of variably controlling a valve-open and/or valve-close timing characteristic of the intake valve. The first variable mechanism varies an oscillating position of an oscillating cam  17  via a transmission mechanism  18  by controlling a rotational position of a control shaft  32  by an electric motor  34 . On the other hand, the second variable mechanism is constructed to change a phase by axially moving a ring gear  43  by way of hydraulic-pressure supply or release to and from first and second hydraulic pressure chambers  49  and  50  so as to produce relative rotation between a timing sprocket  40  and a drive shaft  13 . At the initial state of engine starting, the system operates to drive only the first variable mechanism. Therefore, by using an energy source for the first variable mechanism separated from an energy source for the second variable mechanism, in presence of a failure in one variable mechanism, it is possible to prevent the engine performance from lowering, and also to enhance the engine startability.

TECHNICAL FIELD

The present invention relates to a variable valve operating systemequipped with a first variable mechanism capable of controlling liftcharacteristics of an engine valve such as an intake valve or an exhaustvalve, and a second variable mechanism capable of controlling avalve-open and/or valve-close timing.

BACKGROUND ART

As is generally known, there have been proposed and developed variousvariable valve operating systems that can enhance an engine operatingperformance by enhancing a degree of freedom of valve operatingcharacteristics by way of a combination of a variable lift mechanismcapable of variably adjusting a valve lift characteristic of an intakevalve and a variable valve timing mechanism capable of variablyadjusting valve-open and/or valve-close timing characteristics.

However, in such conventional variable valve operating systems, therewere insufficient studies of a measure to counter in case at least oneof the variable lift mechanism and the variable valve timing mechanismfails with an internal combustion engine in operation, thereby causingseveral problems, that is, mechanical problems such as the interferencebetween at least two of a piston, an intake valve, and an exhaust valve,in particular a problem of remarkably reduced engine performance.

That is to say, in variable valve operating systems disclosed inJapanese Patent Provisional Publication No. 4-287809 and JapaneseUtility Model Provisional Publication No. 3-99860, a variable liftmechanism and a variable valve timing mechanism both use hydraulicpressure as an energy, source. Therefore, in presence of a hydraulicsystem failure, for example, in presence of a failure in anelectromagnetic switching valve disposed in the middle of a hydrauliccircuit or in presence of a failure in a hydraulic equipment included ina hydraulic pressure control system, a malfunction of the variable liftmechanism and a malfunction of the variable valve timing mechanismoccur. In this case, it is difficult to variably control both of avalvelift and avalve timing. As a result of this, it is impossible to producea sufficient engine performance, and also there is a possibility ofremarkably reduced engine performance due to particular engine operatingconditions.

Additionally, even when a hydraulic system failure does not occur, owingto a high viscosity resistance at low oil temperature conditions such asduring an engine start-up period, it is difficult to insure asatisfactory operation of each of the variable lift mechanism and thevariable valve timing mechanism. Due to the high viscosity resistance aswell as a long length of a hydraulic pressure passage, a delay inhydraulic pressure supply from a hydraulic pump to each of the variablelift mechanism and the variable valve timing mechanism occurs. Owing tosuch a system malfunction, there is a technical problem of a reducedcontrol response for each of the variable lift mechanism and thevariable valve timing mechanism to a change in the engine operatingcondition for a while after starting.

As an additional prior art, Japanese Patent Provisional Publication No.8-177434 also discloses a variable valve operating system.

Roughly speaking, the previously-noted variable valve operating systemis equipped with a valve lift control mechanism capable of variablycontrolling a cam lift of an intake valve (serving as an engine valve)or a cam lift of an exhaust valve (serving as an engine valve) byselectively switching from one of a low-speed cam and a high-speed cam(both provided on a camshaft) to the other depending on engine operatingconditions, and a valve timing control mechanism capable of variablycontrolling a valve-open and/or valve-close timing by changing arelative angular phase between the camshaft and a crankshaft dependingon the engine operating conditions.

Additionally, the above-mentioned system is also equipped with a controlmechanism provided to avoid the interference between the intake valveand the exhaust valve by forcibly switching to the low-speed cam bymeans of the valve-lift control mechanism in presence of a failure inthe previously-noted valve timing control mechanism, or by controllingthe valve-open and/or valve-close timing of the engine valve by means ofthe valve timing control mechanism so that a center of operation of avalve lift moves away from a top dead center in presence of a failure inthe previously-noted valve lift control mechanism.

In the aforementioned variable valve operating system, as discussedabove, it is possible to avoid the mechanical problem such as theinterference between the intake valve and the exhaust valve in presenceof a failure in each of the valve lift control mechanism and the valvetiming control mechanism. However, in case either one of the controlmechanisms fails, the system functions to forcibly switch to thelow-speed cam by means of the valve lift control mechanism or to controla valve-open and/or valve-close timing so that the center of operationof a valve lift moves away from the top dead center by virtue of thevalve timing control mechanism. As a matter of course, a valve overlapthrough which both the intake and exhaust valves are open togetherbecomes reduced. Therefore, in a high engine-speed operating range, theengine power output tends to reduce, and thus it is difficult to producea satisfactory engine performance.

DISCLOSURE OF THE INVENTION

In order to accomplish the aforementioned and other objects, accordingto the invention as claimed in claim 1, a variable valve operatingsystem of an internal combustion engine comprises a first variablemechanism capable of variably controlling at least a lift characteristicof an engine valve depending on an engine operating condition, and asecond variable mechanism capable of variably controlling at least avalve-open and/or valve-close timing characteristic of the engine valvedepending on the engine operating condition, characterized in that thefirst variable mechanism is driven by an electric actuator, whereas thesecond variable mechanism is driven by hydraulic pressure of workingfluid, and in that the system inhibits the second variable mechanismfrom being driven and allows only the first variable mechanism to bedriven within an operating range from engine start operation to a timewhen a temperature of the working fluid reaches a predeterminedtemperature value, and the system allows both the first and secondvariable mechanisms to be driven from a time when the temperature of theworking fluid exceeds the predetermined temperature value.

In accordance with the previously-noted invention, a power source of thefirst variable mechanism is different from a power source of the secondvariable mechanism. Except the engine starting period, even in caseeither one of an electric system and a hydraulic system fails, the othervariable mechanism cannot be affected, and thus at least one variablemechanism can be driven normally. Therefore, it is possible to preventthe engine performance from lowering. Furthermore, during the enginestarting period a valve lift is generally adjusted to a small amount,and therefore there is no problem of the previously-discussedinterference even in presence of the system failure.

Additionally, even when both the variable mechanisms are driven,actuation of the second variable mechanism, using hydraulic pressure asa power source, is inhibited during the engine starting period, and isallowed from the time when the temperature of working fluid reaches thepredetermined temperature value. Thus, for a while from the enginestarting, it is possible to enhance various engine performances, such asa startability, owing to high-response lift control of the firstvariable mechanism. Additionally, after a rise in the temperature ofworking fluid, it is possible to largely enhance the engine performancesby way of a combination of both the first and second variablemechanisms.

According to the invention as claimed in claim 2, a valve lift of theengine valve is controlled to a minimum lift substantially correspondingto a substantially zero-lift position during engine crankingcorresponding to an initial stage of engine starting, and then the valvelift of the engine valve is variably controlled so that the valve liftincreases according to a rise in engine speed.

Therefore, during the engine cranking corresponding to the initial stageof engine starting, the hydraulic-pressure driven second variablemechanism cannot be driven, and in lieu thereof the valve lift of theengine valve can be controlled to the minimum lift substantiallycorresponding to a substantially zero-lift position by means of thefirst variable mechanism. Thus, it is possible to provide a smoothengine-speed rise characteristic. Further, the valve lift tends toincrease according to a rise in engine speed, thereby enhancing a gasexchange efficiency, and consequently insuring a good startability.

According to the invention as claimed in claim 3, the above-mentionedfirst variable mechanism comprises a drive shaft having a drive camformed on an outer periphery thereof, and an oscillating cam beingoscillatingly supported on a pivot and acting to open and close theengine valve by way of oscillating motion thereof, and a rocker armrotatably linked at one end to the drive cam and rotatably linked at theother end to the oscillating cam, and a center of the oscillating motionof the rocker arm being variably controlled by means of a control cam.

In accordance with the previously-noted invention, the valve lift can becontinuously varied by rotating the control cam, and additionally it ispossible to set a variable width of valve lift to a greater value. Thesystem of the invention can provide a sufficient engine performance evenin presence of a failure in each of the variable mechanisms as well asin absence of the failure in each of the variable mechanisms.

Additionally, owing to the use of the control cam, a phase of the valvelift uniquely changes according to a change in the valve lift, but thesecond variable mechanism, capable of varying the phase of the driveshaft, is combined with the first variable mechanism, and thus it ispossible to correct the previously-noted unique changes in the phase ofthe valve lift. As a result of this, it is possible to provide asatisfactory engine performance in the absence of the failure in each ofthe variable mechanisms.

According to the invention as claimed in claim 4, the first variablemechanism comprises a drive shaft having a drive cam formed on an outerperiphery thereof, and an axis of the drive cam being eccentric to anaxis of the drive shaft, a link arm rotatably at one end linked to anouter periphery of the drive cam, a rocker arm rotatably linked at oneend to the other end of the link arm, and a center of oscillating motionbeing variably controlled by means of a control cam, an oscillating camacting to open and close the engine valve, a link member mechanicallyand rotatably linking the oscillating cam to the other end of the rockerarm, and an electric actuator controlling a rotational position of thecontrol cam by rotating the control cam by means of a control shaftresponsively to the engine operating condition.

In accordance with the previously-noted invention, by means of the linkmember, it is possible to limit a maximum range of oscillating motion ofthe oscillating cam within a range of oscillating motion of the rockerarm. Thus, even in a high engine speed range, it is possible tocertainly prevent a jumping phenomena, such as excessive oscillation andexcessive jumping motion. Therefore, it is possible to avoid collisionbetween the oscillating cam and the rocker arm, occurring due tomovement of the oscillating cam into and out of contact with the rockerarm, thus preventing occurrence of hammering noise, and also preventingthe accuracy of valve-lift control from lowering. In particular, in thehigh engine speed range, it is possible to stabilize the engineperformance.

According to the invention as claimed in claim 5, a variable valveoperating system of an internal combustion engine comprises a firstvariable mechanism capable of variably controlling at least a liftcharacteristic of an engine valve depending on an engine operatingcondition, a second variable mechanism capable of variably controllingat least a valve-open and/or valve-close timing characteristic of theengine valve depending on the engine operating condition, a positiondetection means for detecting a current actuated position of at leastone of the first and second variable mechanisms, and a control meansbeing responsive to a failure position of the at least one of the firstand second variable mechanisms, detected by the position detectionmeans, for controlling movement of the other variable mechanism within apredetermined range when the at least one of the first and secondvariable mechanisms fails.

In accordance with the previously-noted invention, for example, when thefirst variable mechanism fails in a predetermined engine operatingrange, the position detection means detects a failure position of firstvariable mechanism, and then its informational signal is output into thecontrol means, and thereby the control means can control can control thesecond variable mechanism within the predetermined range capable ofavoiding a mechanical interference between engine valves as much aspossible, responsively to the failure position of the first variablemechanism. Hitherto, the conventional system functioned to forciblyswitch or control to a low-speed cam by means of the first variablemechanism. Therefore, according to the system of the invention, it ispossible to insure a satisfactory engine performance depending on theengine operating condition, as much as possible.

According to the invention as claimed in claim 6, a variable valveoperating system of an internal combustion engine comprises a firstvariable mechanism capable of variably controlling at least a liftcharacteristic of an engine valve depending on an engine operatingcondition, a second variable mechanism capable of variably controllingat least a valve-open and/or valve-close timing characteristic of theengine valve depending on the engine operating condition, a positiondetection means for detecting a current actuated position of the secondvariable mechanism, and a control means being responsive to a failureposition of the second variable mechanism, detected by the positiondetection means, for controlling movement of the first variablemechanism within a predetermined range when the second variablemechanism fails.

According to the invention as claimed in claim 7, a variable valveoperating system of an internal combustion engine comprises a firstvariable mechanism capable of variably controlling at least a liftcharacteristic of an engine valve depending on an engine operatingcondition, a second variable mechanism capable of variably controllingat least a valve-open and/or valve-close timing characteristic of theengine valve depending on the engine operating condition, a positiondetection means for detecting a current actuated position of the firstvariable mechanism, and a control means being responsive to a failureposition of the first variable mechanism, detected by the positiondetection means, for controlling movement of the second variablemechanism within a predetermined range when the first variable mechanismfails.

In accordance with the previously-noted invention as recited in claims 6and 7, in the same manner as the invention as recited in claim 5, whenone of the first and second variable mechanisms fails, by means of thecontrol means it is possible to control the other variable mechanism asmuch as possible, and continuously or intermittently within apredetermined range capable of avoiding a mechanical interferencebetween an engine valve and a piston and a mechanical interferencebetween an intake valve and an exhaust valve, responsively to thefailure position of one variable mechanism. Therefore, it is possible toprevent the engine performance from lowering, while avoiding mechanicalproblems.

According to the invention as claimed in claim 8, the above-mentionedfirst variable mechanism comprises a drive shaft having a drive camformed on an outer periphery thereof, and an oscillating cam beingoscillatingly supported on a pivot and acting to open and close theengine valve by way of oscillating motion thereof, and a rocker armrotatably linked at one end to the drive cam and rotatably linked at theother end to the oscillating cam, and a center of the oscillating motionof the rocker arm being variably controlled by means of a control cam.

In accordance with the previously-noted invention, the valve lift can becontinuously varied by rotating the control cam, and additionally it ispossible to set a variable width of valve lift to a greater value. Thesystem of the invention can provide a sufficient engine performance evenin presence of a failure in each of the variable mechanisms as well asin absence of the failure in each of the variable mechanisms.

Additionally, owing to the use of the control cam, a phase of the valvelift uniquely changes according to a change in the valve lift, but thesecond variable mechanism, capable of varying the phase of the driveshaft, is combined with the first variable mechanism, and thus it ispossible to correct the previously-noted unique changes in the phase ofthe valve lift. As a result of this, it is possible to provide asatisfactory engine performance in the absence of the failure in each ofthe variable mechanisms.

According to the invention as claimed in claim 9, the first variablemechanism comprises a drive shaft having a drive cam formed on an outerperiphery thereof, a link arm rotatably at one end linked to an outerperiphery of the drive cam, a rocker arm rotatably linked at one end tothe other end of the link arm, and a center of oscillating motion beingvariably controlled by means of a control cam, an oscillating cam actingto open and close the engine valve, a link member mechanically androtatably linking the oscillating cam to the other end of the rocker armand limiting a maximum range of oscillating motion of the oscillatingcam within a range of oscillating motion of the rocker arm, and anelectric actuator controlling a rotational position of the control camby rotating the control cam by means of a control shaft responsively tothe engine operating condition.

In accordance with the previously-noted invention, by means of the linkmember, it is possible to limit the maximum range of oscillating motionof the oscillating cam within the range of oscillating motion of therocker arm. Thus, even in a high engine speed range, it is possible tocertainly prevent a jumping phenomena, such as excessive oscillation andexcessive jumping motion. Therefore, it is possible to avoid collisionbetween the oscillating cam and the rocker arm, occurring due tomovement of the oscillating cam into and out of contact with the rockerarm, thus preventing occurrence of hammering noise, and also preventingthe accuracy of valve-lift control from lowering. In particular, in thehigh engine speed range, it is possible to stabilize the engineperformance.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a cross-sectional view illustrating one embodiment of avariable valve operating system of the invention.

FIG. 2 is a cross-sectional view taken along the line A—A shown in FIG.1.

FIG. 3 is a plan view illustrating the first variable mechanism.

FIG. 4 is a cross-sectional view explaining the operation of minimumlift control for the first variable mechanism.

FIG. 5 is a cross-sectional view showing a control process from themaximum lift of the first variable mechanism to the minimum lift.

FIG. 6 is a cross-sectional view explaining the operation of maximumlift control for the first variable mechanism.

FIG. 7 is a characteristic diagram showing valve-lift characteristiccurves and valve-timing characteristic curves of the system of theembodiment.

FIG. 8 is a control flow chart executed by a controller employed in thesystem of the embodiment.

FIG. 9 is a control flow chart executed by the controller employed inthe system of the embodiment.

FIG. 10 is a control flow chart executed by the controller employed inthe system of the embodiment.

FIG. 11 is a control flow chart executed by the controller employed inthe system of the embodiment.

DESCRIPTION OF THE PREFERRED EMBODIMENTS

The present invention will be hereinbelow described in detail inreference to the drawings attached hereto. As shown in FIG. 1, thevariable valve operating system of the invention is exemplified in anintake valve side. Two intake valves 12, 12 per engine cylinder areslidably provided on a cylinder head 11 through valve guides (notshown). A first variable mechanism 1 capable of variably controlling avalve lift of each of the intake valves 12, 12 depending on an engineoperating condition, and a second variable mechanism 2 capable ofvariably controlling a valve-open and/or valve-close timing of each ofthe intake valves 12, 12 are also provided.

As shown in FIGS. 1 through 3, the first variable mechanism 1 includes ahollow drive shaft 13 rotatably supported by means of a bearing 14provided on the upper portion of the cylinder head 11, two drive cams15, 15 constructed by eccentric cams fixedly press-fitted to the driveshaft 13, oscillating cams 17, 17 oscillatingly supported on the driveshaft 13 and capable of opening the intake valves 12, 12 by way ofmovement of the oscillating cams into sliding-contact with respectiveflat upper faces 16 a, 16 a of valve lifters 16, 16 provided at theupper portions of the intake valves 12, 12, a force transmissionmechanism 18 linked between each pair of the drive cams 15 and each ofthe oscillating cams 17, 17, for transmitting a rotational force of thedrive cam 15 as an oscillating force of the associated one of theoscillating cams 17, 17, and a control mechanism 19 capable of variablycontrolling an actuated position of the transmission mechanism 18.

The previously-noted drive shaft 13 is arranged in the longitudinaldirection of the engine. A rotational force is transmitted from anengine crankshaft to the drive shaft through a timing chain (not shown)wound on a timing sprocket 40 of the second variable mechanism 2 (whichwill be fully described later) provided at one end.

As shown in FIG. 1, the bearing 14 is provided on the upper end of thecylinder head 11. The bearing has a main bracket 14 a supporting theupper portion of the drive shaft 13, and a sub bracket 14 b provided atthe upper end of the main bracket 14 a to rotatably support a controlshaft 32 (which will be hereinafter described in detail). Both of thebrackets 14 a and 14 b are firmly secured each other from the upper partby way of a pair of bolts 14 c, 14 c.

As shown in FIGS. 1 through 3, both of the drive cams 15 are formed intoa ring shape. Each of the drive cams includes a cam body 15 a and acylindrical portion 15 b integrally formed on the outside end of the cambody 15 a. The drive cam has an axially-extending drive-shaft insertionhole 15 c formed therein as a through opening. The axis X of the cambody 15 a is offset in the radial direction from the axis Y of the driveshaft 13. Also, each of the drive cams 15 is press-fitted to the driveshaft 13 through the associated drive-shaft insertion hole 15 c outsideof the valve lifters so that the drive cams never interference with therespective valve lifters 16, 16. Outer peripheral surfaces 15 d, 15 d ofboth of the cam bodies 15 a, 15 a have the same cam profile.

As shown in FIG. 2, the above-mentioned oscillating cams 17 are formedinto a substantially U shape in lateral cross section. One end of theoscillating cam is formed as an annular base portion 20. The baseportion is formed with a support hole 20 a into which the drive shaft 13is inserted so that the drive shaft is rotatably supported by thesupport hole. The other end of the oscillating cam is formed as a camnose portion 21. The cam nose portion is formed with a pin hole 21 a.Additionally, the oscillating cam 17 is formed on its lower face with acam surface 22. The oscillating cam is also formed with a basic-circlesurface 22 a of the base portion 20, a circular-arc shaped ramp surface22 b extending from the basic-circle surface 22 a to the cam noseportion 21, and a lift surface 22 c being continuous with the end of theramp surface 22 b. The basic-circle surface 22 a, the ramp surface 22 b,and the lift surface 22 c are designed to be in abutted-contact with apredetermined position of the upper face 16 a of the valve lifter 16,depending on the oscillating position of the oscillating cam 17.

As shown in FIG. 2, the transmission mechanism 18 includes a rocker arm23 arranged above the drive shaft 13, a link arm 24 linking one end 23 aof the rocker arm 23 and the drive cam 15, and a link rod 25 serving asa link member mechanically linking the other end 23 b of the rocker arm23 and the oscillating cam 17.

As shown in FIG. 3, each of the rocker arms 23 is bent or formed into asubstantially crank shape, as seen from a plan view. A cylindrical baseportion 23 c formed at the center of the crank-shaped rocker arm isrotatably supported by a control cam 33 which will be fully describedlater. As seen from FIGS. 2 and 3, the previously-noted one end 23 aprojected from each outside end of the respective base portion 23 c isformed with a pin hole 23 d (a through opening) into which a pin 26 isinserted. The pin is connected to the link arm 24 in such a manner as tobe rotatable relative to the link arm. On the other hand, thepreviously-noted other end 23 b projected from each inside end of therespective base portion 23 c is formed with a pin hole 23 e (a throughopening) into which a pin 27 is inserted. The pin is connected to thelink rod 25 in such a manner as to be rotatable relative to one end 25 aof the link rod 25.

Also, the previously-noted link arm 24 includes a comparativelylarge-diameter, annular base portion 24 a, and a projected end 24 bprojecting from a predetermined position of the outer peripheral surfaceof the base portion 24 a. The base portion 24 a is also formed at itscenter with a loose-fit hole 24 c rotatably fitted onto the outerperipheral surface of the cam body 15 a of the drive cam 15, whereas theprojected end 24 b is formed with a pin hole 24 d (a through opening)into which the pin 26 is inserted so that the pin is rotatably supportedby the pin hole.

Furthermore, as shown in FIG. 2, the link rod 25 is bent or formed intoa substantially L shape of a predetermined length. As shown in FIG. 3,both ends 25 a, 25 b of the link rod are formed with pin insertion holes25 c, 25 d, respectively. The ends of the pins 27 and 28, inserted intothe pin hole 23 e formed in the other end 23 b of the rocker arm 23 andthe pin hole 21 a formed in the cam nose portion 21 of the oscillatingcam 17, are rotatably inserted into the respective pin insertion holes25 c and 25 d.

The above-mentioned link rod 25 is designed to limit the maximum rangeof oscillating motion of the oscillating cam 17 within theoscillating-motion range of the rocker arm 23.

Snap rings 29, 30, and 31 are attached to ends of the respective pins26, 27, and 28, for restricting axial movement of the link arm 24 andaxial movement of the link rod 25.

The previously-noted control mechanism 19 is constructed by the controlshaft 32 arranged in the longitudinal direction of the engine, thecontrol cam 33 fixed onto the outer periphery of the control shaft 32and serving as a fulcrum for oscillating motion of the rocker arm 23,and an electric motor 34 serving as an electric actuator capable ofcontrolling the rotational position of the control shaft 32.

The above-mentioned control shaft 32 is arranged parallel to the driveshaft 13. As discussed above, the control shaft is rotatably supportedbetween a bearing groove of the upper end of the main bracket 14 a ofthe bearing 14 and the sub bracket 14 b. On the other hand, each of thecontrol cams 33 is cylindrical in shape. As shown in FIG. 2, the axis P1of the control cam is deviated from the axis P2 of the control shaft 32by an eccentric distance e.

The previously-noted electric motor 34 is designed to transmit arotational force to the control shaft 32 through a first spur gear 35(mounted on the extremity of a drive shaft 34 a) and a second spur gear36 (mounted on the rear end of the control shaft 32) inmeshed-engagement with each other. The electric motor is driven by acontrol signal from a controller 37 which detects engine operatingconditions.

On the other hand, as shown in FIG. 1, the previously-noted secondvariable mechanism 2 is provided at the extremity of the drive shaft 13.The second variable mechanism is constructed by the timing sprocket 40to which the rotational force is transmitted from the engine crankshaftvia the timing chain (not shown), a sleeve 42 fixedly connected to theextremity of the drive shaft 13 by bolts 41 from the axial direction, acylindrical ring gear 43 interleaved between the timing sprocket 40 andthe sleeve 42, and a hydraulic circuit 44 serving as a drive mechanismcapable of driving the cylindrical ring gear 43 in the axial directionof the drive shaft 13.

The above-mentioned timing sprocket 40 includes a cylindrical body 40 a,a sprocket portion 40 b, and a front cover. The sprocket portion, onwhich the timing chain is wound, is fixed to the rear end of thecylindrical body by means of bolts 45. The front opening of thecylindrical body 40 a is closed by the front cover 40 c. Inner helicalgear 46 is formed on the inner peripheral surface of the cylindricalbody 40 a.

The above-mentioned sleeve 42 is formed at its rear end with a fittinggroove into which the extremity of the drive shaft 13 is fit. A coilspring 47 is disposed in a holding groove of the front end of the sleevefor forwardly biasing the timing sprocket 40 through the front cover 40c. Additionally, outer helical gear 48 is formed on the outer peripheralsurface of the sleeve 42.

The previously-noted ring gear 43 is divided into two parts, namelyfront and rear gear component parts, in a direction perpendicular to theaxial direction. The front and rear gear component parts are biased toeach other by means of pins and springs, so that the two component partsmove towards each other. Additionally, the ring gear is formed at itsinner peripheral surface with an inner helical toothed portion inmeshed-engagement with the outer helical gear 48 and with an outerhelical toothed portion in meshed-engagement with the inner helical gear46. First and second hydraulic pressure chambers 49 and 50 are definedin front and in rear of the ring gear. The ring gear is designed to movein the longitudinal direction while being in sliding-contact with therespective inner and outer gears, by way of hydraulic pressure suppliedrelatively into the first and second hydraulic pressure chambers.Additionally, the ring gear 43 is designed to control the intake valve12 to the maximum timing retard position when the ring gear ispositioned at its maximum forward shifting position at which the ringgear is abutted-engagement with the front cover 40 c, and designed tocontrol the intake valve to the maximum timing advance position with thering gear held at its maximum backward shifting position. Furthermore,the ring gear is designed to be biased to the maximum forward shiftingposition by means of a return spring 51 disposed in the second hydraulicpressure chamber 50 under preload, when there is no supply of hydraulicpressure into the first hydraulic pressure chamber 49.

The hydraulic circuit 44 is constructed by a main gallery 53 connectedto the downstream side of an oil pump 52 communicating an oil pan (notshown), first and second hydraulic pressure passages 54 and 55 branchedat the downstream side of the main gallery 53 and respectively connectedto the first and second hydraulic pressure chambers 49 and 50, asolenoid-actuated fluid-passage directional control valve 56 provided atthe branched position, and a drain passage 57 connected to thefluid-passage directional control valve 56.

The above-mentioned fluid-passage directional control valve 56 isswitched and driven by a control signal from the same controller thatcontrols the electric motor 34 of the first variable mechanism 1.

The controller 37 arithmetically calculates or computes a current engineoperating condition on the basis of input information signals fromvarious sensors, such as an engine-speed indicative signal from a crankangle sensor, an intake-air quantity indicative signal (an engine-loadindicative signal) from an air flow meter, and a signal from an engineoil temperature sensor, and detects and determines the current engineoperating condition. Additionally, the controller receives a signal froma first position detection sensor 58 capable of detecting a currentrotational position of the control shaft 32 and a signal from a secondposition detection sensor 59 capable of detecting a relative rotationalposition of the drive shaft 13 relative to the timing sprocket 40, andoutputs control signals respectively to the electric motor 34 and thefluid-passage directional control valve 56 on the basis of the theseinput information signals from the first and second position detectionsensors. In case that either one of the first and second variablemechanisms 1 and 2 fails and is brought into a locking state, the systemhas a control circuit serving as a control means for variablycontrolling the other variable mechanism continuously within apredetermined range responsively to the locking position (failureposition) of one variable mechanism.

That is to say, the controller 37 determines a desired liftcharacteristic of the intake valve 12, that is, a desired rotationalposition of the control shaft 32, on the basis of the input informationsignals, such as engine speed, engine load, oil temperature, and anelapsed time measured from engine start-up. The controller functions torotate the electric motor 34 in response to a control signalrepresentative of the desired rotational position, and thus the controlcam 33 is rotated and controlled to a predetermined rotational positionvia the control shaft 32. Also, the actual rotational position of thecontrol shaft 32 is monitored by the first position detection sensor 58,so that the control shaft 32 is rotated and brought closer to thedesired phase by way of feedback control.

Concretely, during cranking corresponding to an initial stage of enginestarting and during idling, the control shaft 32 is rotated andcontrolled in one direction by the electric motor 34 driven by thecontrol signal from the controller 37. As shown in FIG. 4, the axis P1of the control cam 33 moves away from the axis P2 of the control shaft32 and is held at a leftward, upward rotational position, and thereforea thick-walled portion 33 a rotates and moves upwards away from thedrive shaft 13. As a result, the rocker arm 23 moves upwards withrespect to the drive shaft 13. Thus, each of the oscillating cams 17 isforcibly pulled up through the link rod 25, and thus rotates in thecounterclockwise direction. Therefore, when rotational movement of thedrive cam 15 pushes up one end 23 a of the rocker arm 23 through thelink arm 24, its lift is transmitted through the link rod 25 to theoscillating cam 17 and to the valve lifter 16. As shown in FIGS. 4 and7, this lift L becomes small. For this reason, gas flow can bestrengthened and thus combustion can be improved. As a result of this,fuel economy can be improved and the engine rotation can be stabilized.

In particular, during engine cranking, as can be seen from FIG. 7, thevalve lift is set to a minimum lift (Lmin) substantially correspondingto a substantially zero-lift position. As described later, it ispossible to smoothly rise the engine speed.

On the other hand, in a high engine-speed range, the electric motor 34is rotated by the control signal from the controller 37, and thus thecontrol shaft 32 is rotated in the other rotational direction by theelectric motor. Such rotation of the control shaft rotates the controlcam 33 in a position indicated in FIGS. 2 and 6, and then thethick-walled portion 33 a is rotated downwards. Thus, the rocker arm 23moves in the direction of the drive shaft 13, (that is, in the downwarddirection), and the other end 23 b pushes down the oscillating cam 17via the link arm 25. As a result, the oscillating cam 17 rotates apredetermined amount to an indicated position (in the clockwisedirection). Therefore, the drive cam 15 rotates so as to push up one end23 a of the rocker arm 23 via the link arm 24. As a result, its lift istransmitted through the link rod 25 to the oscillating cam 17 and to thevalve lifter 16. As shown in FIG. 6, the lift L becomes greatest (themaximum lift Lmax). A change in lift ranging from the minimum lift(Lmin) to the maximum lift (Lmax) is determined depending on therotational position of the control cam 33, as can be seen from thecharacteristic curves (L1-L6) indicated in FIG. 7. In FIG. 7, a liftdenoted by Lmin represents a minimum lift corresponding to a zero lift.However, assuming that the control shaft is further rotated in the onedirection as previously discussed, the value Lmin can be set at zero.

On the other hand, as regards a side of the fluid-passage directionalcontrol valve 56, in the same manner as discussed, a desired timingadvancement of the intake valve 12 is determined on the basis of theinformation signals from the respective sensors. The fluid-passagedirectional control valve 56 is switched in response to a control signalrepresentative of the desired timing advancement in a manner so as tointercommunicate the first hydraulic pressure passage 54 and the maingallery 53 for a predetermined time interval, and also tointercommunicate the second hydraulic pressure passage 55 and the drainpassage 57 for a predetermined time interval. Thereby, the relativerotational position between the timing sprocket 40 and the drive shaft13 is changed through the ring gear 43, so that the valve timing of theintake valve is controlled to a timing-advance position. Also, theactual relative rotational position of the drive shaft 13 is monitoredby the second position detection sensor 59, so that the drive shaft isrotated and brought closer to the desired relative rotational position,that is, the desired timing advancement by way of feedback control.

Concretely, until the predetermined time period has expired from enginestarting, that is, until the oil temperature has reached thepredetermined temperature value To, hydraulic pressure is supplied intoonly the second hydraulic pressure chamber 50 by means of thefluid-passage directional control valve 56, and hydraulic pressure isnot supplied into the first hydraulic pressure chamber 49. Therefore, ascan be seen from FIG. 1, the ring gear 43 is held at the maximum forwardshifting position by virtue of the bias of the return spring 51, andthen the drive shaft 13 is held at the rotational position correspondingto the maximum timing retardation. After this, as soon as the oiltemperature exceeds the predetermined temperature value To, thefluid-passage directional control valve 56 is driven in response to thecontrol signal from the controller 37 depending on the engine operatingcondition, with the result that the first hydraulic pressure passage 54is communicated with the main gallery 53, and also a time intervalduring which the second hydraulic pressure passage 55 and the drainpassage 57 are communicated with each other continuously varies. As aresult of this, the ring gear 43 moves from its maximum forward shiftingposition to its maximum backward shifting position. Therefore, thevalve-open and/or valve-close timing of the intake valve 12 can bevariably controlled continuously from the maximum timing-retard stateindicated by the solid line shown in FIG. 7 to the maximumtiming-advance state indicated by the broken line shown in FIG. 7.

In a state wherein the previously-noted intake valve 12 is controlled toits maximum lift by means of the first variable mechanism 1, andsimultaneously controlled to its maximum timing-retard position by meansof the second variable mechanism 2, the intake valve is arranged andconstructed so that there is no interference with the piston disposed inthe cylinder and with the exhaust valve opposing thereto.

Hereunder explained in detail in reference to the flow chart shown inFIGS. 8 and 9, is the concrete control necessary to drive each of thefirst and second variable mechanisms 1 and 2.

First of all, considering the oil temperature after engine starting, asshown in FIG. 8, at step S1, a check is made to determine whether anelapsed time counted from the engine starting exceeds the predeterminedtime period to based on a counted value of a timer. When exceeding thepredetermined time period, the routine proceeds to step S2 in which, onthe basis of information from the oil temperature sensor, a check ismade to determine whether the current oil temperature exceeds thepredetermined temperature value To. When exceeding the predeterminedtemperature value, the routine proceeds to step S3 in which both of thefirst and second variable mechanisms 1 and 2 are driven. Conversely,when the elapsed time does not exceed the predetermined time period toat step S1 or when the oil temperature is below the predeterminedtemperature value To at step S2, the routine proceeds to step S4 inwhich the control for the variable mechanisms is made so that only thefirst variable mechanism 1 is driven and the second variable mechanism 2is not driven.

Therefore, during engine starting at low temperatures, only thevalve-lift control executed by the first variable mechanism 1 is made,and the valve-timing control executed by the second variable mechanism 2is not made, and as a result the intake valve 12 is held at the maximumtiming-retard state. Accordingly, in such an operating range, there isno problem of the variable mechanism failure which may occur owing tothe hydraulic energy source. Additionally, it is possible to enhanceengine startability by virtue of the valve-lift control, and to enhancethe engine performance. Also, after the oil temperature has risen, thesecond variable mechanism 2 can be driven, thus remarkably largelyenhancing the engine performance.

If a failure in the electrical system of the first variable mechanism 1occurs, there is a possibility that the control shaft 32 is affected bya reaction force created in the valve operating system, and thus anundesirable phase change occurs. However, the second variable mechanism2 is kept at the maximum timing-retard position. Thus, even when theintake valve has reached the maximum lift owing to the phase change ofthe control shaft, there is no mechanical problem such as a mechanicalinterference between the intake valve 12 and the piston.

Hereinbelow described in reference to FIG. 9 is the control routine forthe first variable mechanism 1. First, at step S11, the ignition switchis turned on. Thereafter, at step S12, the first variable mechanism 1 iscontrolled to the minimum lift (substantially corresponding to asubstantially zero-lift position). Subsequently to the above, at stepS13, the starter switch is turned on, and then engine cranking isinitiated. After this, at step S14, the valve lift is controlled bymeans of the first variable mechanism 1, so that the lift increases upto the characteristic curve indicated by the solid line L3 shown in FIG.7 according to an increase in engine speed (cranking speed).

Thereafter, at step S15, a check is made to determine whether thecurrent oil temperature detected by the oil temperature sensor is higherthan the predetermined temperature value (T1). When the current oiltemperature is higher than the predetermined temperature value, theroutine proceeds to step S16 in which the variable lift control suitablefor the current engine operating condition is executed by the firstvariable mechanism. Conversely, when the current oil temperature is lessthan or equal to the predetermined temperature value T1, the routineproceeds to step S17 in which the lift control is executed by the firstvariable mechanism 1 so that the valve lift is fixed to thepreviously-discussed characteristic curve L3.

In this manner, at the initial state of engine starting (just aftercranking), the valve lift is controlled to the minimum lift through stepS12, and therefore the friction of the valve operating system becomessmaller. Thus, it is possible to quickly rise the engine speed.

Additionally, due to the lift-increase control achieved through stepS14, the gas exchange efficiency of the air-fuel mixture can beenhanced, and thus the engine torque can rise quickly. Conjointly withthe quick engine-speed rise as previously discussed, it is possible tolargely improve the engine startability.

Furthermore, in case that the oil temperature is below the predeterminedtemperature value T1, the valve lift is fixed to a relatively low liftcorresponding to the lift characteristic L3 through step S17. Thus, theflow velocity of the air-fuel mixture gas flow passing through theintake valve 12 can be increased, thereby producing a strong gas flow inthe engine cylinder. As a result, it is possible to improve combustionduring cold engine starting, and thus improving fuel economy andenhancing exhaust emission control performance.

Additionally, the first variable mechanism 1 of the embodiment exhibitsthe valve lift characteristics shown in FIG. 7. From a study of a phase(a valve lift phase) of the drive shaft 13 at which the valve liftbecomes the maximum valve lift, the system exhibits a unique valve-liftphase characteristic according to which the valve timing advances littleby little as the valve lift reduces from the maximum lift Lmax, andthereafter when the valve lift further reduces towards the minimum liftLmin the valve timing conversely retards midway. This is because, asshown in FIG. 6, a timing of the maximum valve lift corresponds to amoment that the radius vector R1 of the eccentric circle of the drivecam 15 and the line segment R2 interconnecting the axis X of the drivecam 15 and the pivot Z of the projected end 24 b of the link arm 24 arealigned with each other. At this time, the direction of the radiusvector R1 is offset from the vertical line Q of the cylinder head 11 byan angle θ in the valve-timing advance direction.

Next, as viewed from a case where the control shaft 32 rotates in theclockwise direction (viewing the drawing), and then the radius vector R1of the drive cam 15 and the link arm 24 are aligned with each other.That is, at this time, the angle θ gradually increases as the controlshaft 32 rotates in the clockwise direction, and then becomes maximumwhen the radius vector R3 of the rocker arm 23 and the radius vector eof the control cam 33 are aligned with each other (see FIG. 5), andconversely reduces (see FIG. 7) when the control shaft 32 furtherrotates in the clockwise direction as shown in FIG. 4. For the reasonsset forth above, as a consequence, the valve-lift phase varies accordingto the unique characteristic.

Additionally, the controller 37 performs control routines shown in FIGS.10 and 11 by means of the control circuit, when the first variablemechanism or the second variable mechanism fails in the operating rangein which the first and second variable mechanisms are both variablycontrolled.

According to the control routine shown in FIG. 11, first of all, at stepS31, information from each of the sensors is read. At step S32, theactual rotational position (corresponding to a lift) of the controlshaft 32 is read on the basis of information from the first positiondetection sensor 58. Thereafter, at step S33, a check is made todetermine whether the first variable mechanism 1 fails, by way ofcomparison between the actual rotational position and the desiredrotational position. When the first variable mechanism failure isdetermined, the routine proceeds to step S34 in which a controlledposition for the second variable mechanism, that is, a control range (atiming advancement) in which there is no interference between the intakevalve 12 and the piston and there is no interference between the intakevalve 12 and the exhaust valve, is arithmetically calculated.Furthermore, at step S35, the second variable mechanism 2 iscontinuously controlled within the predetermined control range.

That is, in the case that the first variable mechanism 1 fails duringthe maximum lift (Lmax) control, the second variable mechanism 2 iscontinuously controlled in the vicinity of the maximum timing-retardposition in order to avoid the interference between both of the enginevalves. Also, in the case that the first variable mechanism fails in thesmall lift range (Lmin through L1), the second variable mechanism 2 iscontinuously controlled in a wide range extending from the maximumtiming retard position to the maximum timing advance position. As aresult of this, it is possible to suppress the engine performance frombeing deteriorated. Furthermore, in the case that the first variablemechanism fails in the middle lift (L3) range, the second variablemechanism is continuously controlled in a range extending from themaximum timing retard position to the middle phase.

As discussed above, the second variable mechanism 2 can be continuouslycontrolled within the control range capable of avoiding the interferencebetween each of the engine valves and the piston, and thus it ispossible to prevent the engine performance from lowering.

According to the control routine shown in FIG. 10, at step S21,information from each of the sensors is read. Thereafter, at step S22,the actual relative rotational position (corresponding to a timingadvancement) of the drive shaft 13 is read on the basis of informationfrom the second position detection sensor 59. Thereafter, at step S23, acheck is made to determine whether the second variable mechanism 2fails, by way of comparison between the actual relative rotationalposition and the desired relative rotational position.

When the second variable mechanism failure is determined, the routineproceeds to step S24 in which a controlled position for the firstvariable mechanism, that is, a control range (a lift) in which there isno interference between the intake valve 12 and the piston and there isno interference between the intake valve 12 and the exhaust valve, isarithmetically calculated. Furthermore, at step S25, the first variablemechanism 1 is continuously controlled within the predetermined controlrange.

That is, in the case that the second variable mechanism 2 fails duringthe maximum timing advance control, the first variable mechanism 1 iscontinuously controlled in the small lift range (Lmin through L1) shownin FIG. 7 in order to avoid the interference. In the case that thesecond variable mechanism fails in the maximum timing retard side, thereis no problem of the interference and thus the first variable mechanismis continuously controlled in the entire range extending from theminimum lift to the maximum lift. Furthermore, in the case that thesecond variable mechanism fails within the middle phase, the firstvariable mechanism is continuously controlled in a range extending fromthe minimum lift to the middle lift L3.

As discussed above, even when the second variable mechanism 2 fails, thefirst variable mechanism 1 can be continuously controlled within thecontrol range capable of avoiding the interference between the intakevalve 12 and the piston, and thus it is possible to suppress the engineperformance from lowering, as much as possible. Furthermore, when thesecond variable mechanism fails in the middle phase, the first variablemechanism can be continuously controlled within the control rangeextending from the minimum lift to the middle lift L3.

As discussed above, even when the second variable mechanism 2 fails, thefirst variable mechanism 1 can be continuously controlled within thecontrol range capable of avoiding the interference between the intakevalve 12 and the piston, and thus it is possible to suppress the engineperformance from lowering, as much as possible. Also, it is possible toprovide the same effects by continuously controlling the first variablemechanism in a multi-stage fashion. In this case, the control can besimplified.

Additionally, according to the system of the present embodiment, theoscillating cam 17 is linked to the rocker arm 23 via the link rod 25,and thus the maximum range of oscillating motion of the oscillating cam17 can be regulated within the oscillating motion range of the rockerarm 23 by means of the link rod 25. Therefore, even in the highengine-speed range, it is possible to certainly prevent a jumpingphenomena, such as excessive oscillation and excessive jumping motion ofthe oscillating cam 17. Therefore, it is possible to avoid collisionbetween the oscillating cam 17 and the rocker arm 23, occurring due tomovement of the oscillating cam into and out of contact with the rockerarm, thus preventing occurrence of hammering noise, and also preventingthe accuracy of valve-lift control from lowering. In particular, in thehigh engine speed range, it is possible to stabilize the engineperformance.

Moreover, although in the system of the present embodiment thevalve-lift phase uniquely changes according to a change in valve lift,it is possible to correct the unique changes in the valve-lift phase bycombining the first variable mechanism 1 with the second variablemechanism 2 capable of varying the rotational phase of the drive shaft13. That is to say, under the engine operating condition kept is a highengine-speed range or a high engine-load range, if the valve lift iscontrolled to a large valve lift by means off the first variablemechanism 1 and additionally the valve-lift phase is controlled toapproach to the top dead center, a valve overlap can be adjusted to agreater value. Thus, it is possible to scavenge the residual gasprevailing in the engine cylinder by synchronizing a negative pressurewave of exhaust pulse with the greater valve overlap period, therebyenhancing the charging efficiency of fresh air. As a consequence, it ispossible to greatly enhance the engine power output.

While the foregoing is a description of the preferred embodimentscarried out the invention, it will be understood that the invention isnot limited to the particular embodiments shown and described herein,but that various changes and modifications may be made without departingfrom the scope or spirit of this invention as defined by the followingclaims.

INDUSTRIAL APPLICABILITY

As discussed above, the present invention is not limited to theembodiments shown and described herein, but it will be appreciated thatthe concept of the invention can be applied to the exhaust valve side.In the same manner as the intake valve 12, it is possible to reduce afriction of an operated engine valve by controlling the first variablemechanism 1 to the minimum lift substantially corresponding to asubstantially zero-lift position at the initial stage of enginestarting. Thus, it is possible to ensure a smooth engine-speed risecharacteristic. Furthermore, it is possible to enhance the gas exchangeefficiency by variably controlling the valve lift according to theincrease in engine speed, thus ensuring a good startability. Asdiscussed above, it is possible to provide the same effects as theintake valve side.

In this manner, in the case that the concept of the invention is appliedto the exhaust valve side, if either one of the variable mechanismsfails, it is possible to properly control the other variable mechanismin the same manner as the intake valve side. As a matter of course, itis possible to prevent the engine performance from lowering, whileavoiding the mechanical problem as described previously.

Additionally, in the system of the invention as recited in claims 5through 8, the energy source for each of the variable mechanisms is notlimited to hydraulic pressure or to electrical power. In lieu thereof,any power source may be used. Also, the concept of the invention may beapplied to a system in which both of variable mechanisms are drivenelectrically or hydraulically.

What is claimed is:
 1. A variable valve operating system of an internalcombustion engine comprising: a first variable mechanism capable ofvariably controlling at least a lift characteristic of an engine valvedepending on an engine operating condition; and a second variablemechanism capable of variably controlling at least a valve-open and/orvalve-close timing characteristic of the engine valve depending on theengine operating condition, characterized in that the first variablemechanism is driven by an electric actuator, whereas the second variablemechanism is driven by hydraulic pressure of working fluid, and in thatthe system inhibits the second variable mechanism from being driven andallows only the first variable mechanism to be driven within anoperating range from engine start operation to a time when a temperatureof the working fluid reaches a predetermined temperature value, and thesystem allows both the first and second variable mechanisms to be drivenfrom a time when the temperature of the working fluid exceeds thepredetermined temperature value.
 2. The variable valve operating systemof the internal combustion engine as claimed in claim 1, characterizedin that a valve lift of the engine valve is controlled to a minimum liftsubstantially corresponding to a substantially zero-lift position duringengine cranking corresponding to an initial stage of engine starting bythe first variable mechanism, and then the valve lift of the enginevalve is variably controlled so that the valve lift increases accordingto a rise in engine speed.
 3. The variable valve operating system of theinternal combustion engine as claimed in claim 1, characterized in thatthe first variable mechanism comprises a drive shaft having a drive camformed on an outer periphery thereof, and an oscillating cam beingoscillatingly supported on a pivot and acting to open and close theengine valve by way of oscillating motion thereof, and a rocker armrotatably linked at one end to the drive cam and rotatably linked at theother end to the oscillating cam, and a center of the oscillating motionof the rocker arm being variably controlled by means of a control cam.4. The variable valve operating system of the internal combustion engineas claimed in claim 1, characterized in that the first variablemechanism comprises a drive shaft having a drive cam formed on an outerperiphery thereof, a link arm rotatably at one end linked to an outerperiphery of the drive cam, a rocker arm rotatably linked at one end tothe other end of the link arm, and a center of oscillating motion beingvariably controlled by means of a control cam, an oscillating cam actingto open and close the engine valve, a link member mechanically androtatably linking the oscillating cam to the other end of the rockerarm, and an electric actuator controlling a rotational position of thecontrol cam by rotating the control cam by means of a control shaftresponsively to the engine operating condition.